All gear infinitely variable transmission

ABSTRACT

An infinitely-variable transmission consists of a rotational input member, a rotational output member, first and second non-circular driving gears coupled to one of the rotational members, and a plurality of variable velocity-ratio gear assemblies disposed about the one rotational member. Each gear assembly comprises an intermediate shaft including first and second non-circular driven gears respectively meshing with the first and second non-circular driving gears. Each gear assembly also comprises a multi-directional coupling associated with the pair of first driven/driving gears and the pair of second driven/driving gears, and an actuator associated with the coupling for coupling the first and second variable-ratio gear pairs to the rotational members. Preferably, the variable-ratio gear pairs for each gear assembly are coupled to the rotational members over an angular duration when the velocity ratios of both the variable-ratio gear pairs of the respective gear assembly are changing linearly so as to provide a uniform velocity ratio over that angular duration. Further, preferably the occurrences of these angular durations for each of the gear assemblies overlap or are at least coterminous so as to provide a continuously uniform velocity ratio between the rotational members. The transmission also includes a phase angle variator associated with the variable-ratio gear pairs for varying the rotational angular displacement between the first gear pairs and the second gear pairs of each gear assembly so as to vary the velocity ratio of the transmission as needed.

FIELD OF THE INVENTION

The present invention relates to gear systems having variable velocityratios. In particular, the present invention relates to all-geartransmissions whose velocity ratios may be varied continuously over thevelocity ratio continuum.

BACKGROUND OF THE INVENTION

The conventional transmission incorporates a number of planetary gearswhich are selectively coupled between the input and output shafts forchanging the velocity ratio of the transmission. However, theconventional transmission results in an inefficient transfer of torquebetween the input and output shafts since the prime mover must beuncoupled from the transmission output shaft while the planetary gearsare switched. Further, since engine speed must vary in each gear ratioto affect speed change of the output shaft, the efficiency of the enginecannot be maximized for any particular operating condition. Therefore,many attempts have been made to provide a transmission whose velocityratios are infinitely variable over the velocity ratio continuum.

For instance, Beschkine (U.S. Pat. No. 2,239,313) teaches a gear systemincorporating continuously-meshing non-circular gears. As shown in FIG.3 of the patent, the gear system comprises a driving shaft P including aplurality of non-circular gears 1, 2, 3, 4, and a driven shaft Rparallel to the driven shaft P including a plurality of non-circulargears, 1′, 2′, 3′, 4′ meshing with the gears 1, 2, 3, 4. The drivengears 1′, 2′, 3′, 4′ are coupled successively to the driven shaft R byelectromagnetic clutches for a respective portion of the interval ofrevolution of the driven shaft R so that the velocity ratio of the gearsystem is dependent upon the gear ratios of the gears 1-1′, 2-2′, 3-3′,4-4′ over their respective coupling intervals. Consequently, thevelocity ratio of the gear system is varied by simply changing theangular position of the gears 1′, 2′, 3′, 4′ during their couplingintervals. However, as the clutches must be activated each revolution ofthe driven shaft R, the clutches must be relatively small to be used forhigh speed applications. thereby limiting the torque which can becarried by the gear system.

Kerr(Canadian Patents 990,103; 1,000,526: 1,031,190; U.S. Pat. Nos.3.919.895: 4,055,091) teaches variable output transmission incorporatingsquare-wave generators for facilitating changes to the velocity ratio.Each transmission comprises a pair of non-circular driving gears coupledto an input shaft, and a pair of non-circular driven gears whichcontinuously mesh with the driven gears. The velocity ratio profile ofeach non-circular gear pair resembles a triangular or saw-tooth wave.The rotational outputs of the two non-circular gear pairs arecombined-through a differential to provide a differential output havinga square-wave velocity ratio profile. The differential outputs of anumber of such differentials are combined together through one-wayoverrunning clutches to a provide a velocity ratio which is infinitelyvariable in accordance with the relative angular displacement of thedriving gears. The variable output transmissions taught by Kerrrepresented a significant advance over the prior art. However.overrunning clutches can only transfer energy in a single direction.thereby precluding engine drag. Further. as the variable outputtransmissions only amplified the positive or negative periods of thesquare-wave velocity ratio profile. the efficiency and maximum kinematicrange of the transmissions was limited.

Takahara (U.S. Pat. No. 4,944,718) teaches an angular velocitymodulating device which, as shown in FIGS. 1 to 3 of the patentcomprises a first rotatable shaft 24 rotatably coupled to a stationaryfirst frame 61: non-circular internal gears 11 mounted on the firstshaft 24: a second parallel shaft 34 rotatably mounted on a rotatablesecond frame 62: second non-circular gears 21 meshing with the firstnon-circular gears 11 and fixed on the parallel second shaft 24: thirdnon-circular gears 31 meshing with the first non-circular gears 11 andmounted on the parallel third shaft 34 through an overrunning clutch 37;an input shaft 44 including a circular gear 27 for rotating the secondshaft 24: and an output shaft 54 including a circular gear 54 driven bythe third shaft 34. Since the velocity ratio of the device is varied bychanging the angular displacement of the second frame 62 relative to thefirst frame 61, rapid changes in velocity ratio would be difficult toattain since the angular displacement of the second frame 62 could onlybe changed by also moving the second shaft 34 and the mass of theaccompanying non-circular gears 34. Further, as discussed above. theoverrunning clutches preclude engine drag and reduce the efficiency andmaximum kinematic range of the device.

Recently, Pires (U.S. Pat. Nos. 5,226,859; 5,334,115) disclosed aninfinitely variable transmission which eliminates the need foroverrunning clutches. As shown in FIGS. 2, 3 and 4 of the '859 patent,the transmission comprises an input shaft 5, a planetary rotor 29connected to the input shaft 5 through a front plate 5′. a first pair ofcrank arms 8 a, 8 c rotatably coupled to the planetary rotor 29, asecond pair of crank arms 9 b. 9 d rotatably coupled to the planetaryrotor 29, and an index plate 7 which incorporates slots for receiving anend of the crank arms. The index plate 7 is supported on an index slide6 which allows the index plate to move laterally of the shaft 5. Thetransmission also includes four planar differential gear sets, each setcomprising an internally-toothed ring gear 12 coupled to one of thecrank arms, a pair of pinions 14 meshing with the ring gear 12, and asun gear 15 meshing with the pinions 14. The sun gear 15 of eachdifferential gear set is connected to a reaction gear 16 which mesheswith an internally-toothed stationary commutator gear 28.

In operation, when the input shaft 5 rotates, the planetary rotor 29 isforced to rotate, causing the crank arms to drive the index plate 7about its own axis of rotation, as defined by the index slide 6. If theindex plate 7 is eccentric to the axis of the planetary rotor 29, thecrank arms oscillate about their own axes while orbiting the planetaryrotor 29. The amplitude of oscillations is a function of theeccentricity of the index plate 7. The rotational oscillations aredelivered to the differential gear sets by the ring gears 12. As shownin FIG. 7 of the patent, the commutator gear 28 includes teeth onlyaround half of the inner circumference of the gear, so that the reactiongears 16 rotate freely one half of a rotational cycle of the input shaft5. Consequently, when the “desired” polarity of oscillation is presentat the crank arm, the commutator gear 28 provides a supplementalrotational input to the differential, whereas when the “undesired”polarity of oscillation is present, the commutator gear 28 prevents theoscillation from reducing the output of the transmission.

Although the transmission taught by Pires addresses the problems imposedby overrunning clutches on efficiency and kinematic range, thetransmission is quite complex. Further, it is believed that theoscillating crank arms will produce undesirable fluctuations in thevelocity ratio of the transmission. Accordingly, there remains a needfor an infinitely-variable all gear transmission which has an enhancedkinematic range, is capable of making rapid changes in velocity ratios.and can take advantage of engine drag.

Further. conventional non-circular gears employ standard involute-shapedgear teeth. Although the involute-shaped gear teeth are acceptable foruse with circular gears. involute-shaped gear teeth when used onnon-circular gears cause the contact ratio between the gears tocontinuously vary. These variations in contact ratio cause excessivegear noise. Also. the contact ratio of involute-shaped teeth even oncircular gears rarely reaches 2.0. Consequently, the load which can becarried by the gears is limited. Although the contact ratio may beincreased by twisting the gear teeth, twisted gear teeth produce pointcontact which creates Hertzian stress. Accordingly, there also remains aneed for non-circular gears having a constant contact ratio which ispreferably at least 2.0.

SUMMARY OF THE INVENTION

According to the invention, there is provided an infinitely-variabletransmission which addresses the deficiencies of the prior art.

The infinitely-variable transmission, according to the presentinvention, comprises a rotational input member, a rotational outputmember, a pair of variable velocity-ratio gear sets, a multi-directionalcoupling associated with the gear sets. and an actuator associated withthe coupling. The coupling couples the gear sets to the rotationalmembers over a common angular period so as to provide a uniform velocityratio between the rotational members over the angular period. Thetransmission also includes a phase angle variator associated with atleast one of the gear sets for varying a rotational angular displacementbetween the gear sets. Consequently, the velocity ratio of thetransmission can be varied while maintaining the velocity ratio uniformover the angular period.

In a preferred embodiment of the invention, the transmission comprises afirst and second non-circular driving gears coupled to one of therotational members, and a plurality of variable velocity-ratio gearassemblies disposed about the one rotational member. The gear assembliesare coupled to the non-circular gears and the other of the rotationalmembers for providing a uniform velocity ratio between the rotationalmembers. Each gear assembly comprises an intermediate shaft includingfirst and second non-circular driven gears meshing respectively with thefirst and second non-circular driving gears. Preferably, the velocityratio of each variable-ratio pair of first driven/driving gears and thevelocity ratio of each variable-ratio pair of second driven/drivinggears includes a constant acceleration portion.

Each gear assembly also comprises a multi-directional couplingassociated with the respective first and second variable-ratio gearpairs, and an actuator associated with the coupling for coupling therespective variable-ratio gear pairs to the rotational members.Preferably, the variable-ratio gears of each gear assembly are coupledto the rotational members over an angular period when the velocityratios of both the first and second variable-ratio gear pairs of therespective gear assembly are changing linearly so as to provide auniform velocity ratio over that angular period. Outside this period, atleast one of the respective variable-ratio gear pairs is uncoupled fromthe rotational members. Further, preferably the occurrences of theseangular durations for each of the gear assemblies overlap or are atleast coterminous so as to provide a continuously uniform velocity ratiobetween the rotational members.

The transmission also includes a phase angle variator associated withthe variable-ratio gear pairs for varying the rotational angulardisplacement between the first gear pairs and the second gear pairs ofeach gear assembly so as to vary the velocity ratio of the transmissionas needed.

Variable-ratio gear pairs having a constant contact ratio, and a methodfor defining teeth flanks for such gears for providing a constantcontact ratio are also disclosed. In accordance with the method, toothflanks for the constant contact ratio variable-ratio gears are definedby first determining a pitch locus for one of the non-circular gears.Then the pitch locus is segmented into pitch locus portions. Aneffective pitch circle locus for the pitch locus portions is thendetermined by projecting the pitch locus portions onto a centre linejoining centres of the non-circular gears. An effective generatingcircle locus for the effective pitch circle locus is then determined, inaccordance with a desired pressure angle between the non-circular gears.Finally, a locus of congruency for the gears is determined from theeffective generating circle locus.

BRIEF DESCRIPTION OF THE DRAWINGS

The preferred embodiments of the invention will now be described, by wayof example only, with reference to the drawings, in which:

FIGS. 1, and 1 a to 1 f depict certain prior art infinitely-variabletransmissions;

FIGS. 2, and 2 a to 2 h depict certain infinitely-variable transmissionsaccording to the present invention, showing the variable velocity-ratiogear set. the multi-directional coupling, and the shadow cam discactuator;

FIGS. 3 a to 3 d depict certain variations of the multi-directionalcoupling;

FIGS. 4 a to 4 c depict a variation of the shadow cam disc actuator;

FIGS. 5 a, 5 b depict a variation of the shadow cam disc actuator,incorporating a rocker arm;

FIGS. 6, 6 a, 6 b depict further variations of the shadow cam discactuator,

FIGS. 7, and 7 a to 7 c depict further variations of the shadow cam discactuator, incorporating roller ball followers;

FIGS. 8, 8 a, 9, 9 a, 10 and 10 a depict further variations of theshadow cam disc actuator shown in FIGS. 7, and 7 a to 7 c;

FIGS. 11, 11 a to 11 b depict the phase angle variator for use with theinfinitely-variable-transmissions of the present invention;

FIGS. 12, 13, 13 a, 13 b, 14, 14 a, 14 b depict certain variations ofthe phase angle variator,

FIGS. 15, and 15 a to 15 i depict a preferred tooth flank shape for useas a part of the variable velocity-ratio gear set; and

FIGS. 16, 16 a, 16 b, 17 depict certain preferred implementations of theinfinitely-variable transmissions according to the present invention.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

To assist in understanding the invention, certain prior artinfinitely-variable transmissions will be described first, withreference to FIGS. 1 a to 1 f, followed by a discussion of oneembodiment of the invention beginning with FIG. 2. FIGS. 1, 1 a shows apartial layout of the components of a prior art infinitely-variabletransmission. The transmission comprises an output gear 1 coupled to anoutput shaft 1.1, a first input shaft 3.1 including a first non-circulardriver gear 3, and a second input shaft 12.1 rotatably coupled to thefirst input shaft 3.1 and including a second non-circular driver gear12. The second non-circular gear 12 may be rotated angularly withrespect to the first non-circular gear 1 by rotating the second inputshaft 12.1 relative to the first input shaft 3.1 through a phase anglevariator (not shown).

The prior art infinitely-variable transmission also includes fouridentical variable-ratio gear assemblies disposed at equal angularintervals around the input shaft 3.1 and the output shaft 1.1. Forconvenience. only one variable-ratio gear assembly is shown. Eachvariable-ratio gear assembly comprises an intermediate shaft 13. a firstnon-circular driven gear 4 rotatably disposed around the intermediateshaft 13 and meshing with the first non-circular driver gear 3. a secondnon-circular driven gear 11 rotatably disposed around the intermediateshaft 13 and meshing with the second non-circular driver gear 12, and adifferential coupled to the first and second driven gears 4, 11 forcombining the torque delivered to the output shaft 1.1 from the firstand second gear sets. The differential includes a first bevel gear 8, asecond bevel gear 10, and a cage 9 including pinions 9.1 meshing withthe first and second bevel gears 8, 10. The cage 9 is coupled to theintermediate shaft 13, to which is splined a take-off gear 2. Thetake-off gear 2 itself is coupled to the output gear 1 and, therefore,the output shaft 1.1. The second bevel gear 10 is splined to be seconddriven gear I 1t whereas the first bevel gear 8 is coupled to the firstdriven gear 4 through a one-way overrunning clutch. Consequently, thetransmission shown in FIGS. 1, 1 a is referred to as a “bevel-bevel”transmission.

The overrunning clutch comprises a tubular outer clutch slipper 7. and atubular inner clutch body 5 provided within the clutch slipper 7 andbeing splined to the first non-circular driven gear 4. As shown in thebreakout diagram of FIG. 1, the inner tubular surface of the clutchslipper 7 and the outer tubular surface of the clutch body 5 define aplurality of spiral channels in which are retained a plurality of clutchrollers 6. Consequently, the overrunning clutch will lock in onedirection and will free-wheel in the opposite direction.

The operation of the prior art infinitely-variable transmission shown inFIG. 1 can be understood by referring to the diagrams shown in FIG. 1 b.Diagram 1 of FIG. 1 b shows the variation in velocity ratio W₄ of thepair of first non-circular gears 3,4, the variation in velocity ratioW₁₁ of the pair of second non-circular gears 12, 11, and the variationin velocity ratio R through the transmission over one revolution of theinputs shaft 3.1, 12.1 when the first non-circular gears 3,4 are inphase with the second non-circular gears 12, 11. As will be apparent,the velocity ratio of the first non-circular gears 3,4 increaseslinearly over the first 220° of input shaft rotation, and then decreasesnon-linearly over the subsequent 140° of input shaft rotation.

The velocity ratio R is given by the equation R=(W₄+W₁₁)/2.Consequently, the velocity ratio of the second non-circular gears 12, 11decreases linearly over the first 220° of input shaft rotation at thesame rate as the velocity ratio of the first non-circular gears 3.4increase during this interval to provide a constant velocity ratio R.The velocity ratio of the second non-circular gears 12, 11 increasesover the subsequent 140° of input shaft rotation.

As shown in Diagram 1, the velocity ratio R does remain constant at 1.0over the first 220° of input shaft rotation, as expected. Thereafter,the intermediate shaft 13 overruns the first bevel gear 8, causing thefirst and second non-circular gear pairs to become uncoupled from theinput and output shafts. Consequently, after this point and until theinput shafts return to their starting position, the variable-ratio gearassembly does not transfer any torque to the output shaft 1.1. However,each of the three other variable-ratio gear assemblies (not shown) willbegin operating in sequence after each 90° of input shaft rotation.Consequently, the velocity ratio R of the transmission remains constantat 1.0 over the entire 360° revolution of the input shafts 3.1, 12.1.

Diagram 2 shows the variation in the velocity ratio W₄ of the pair offirst non-circular gears 3.4, the variation in the velocity ratio W₁₁ ofthe pair of second non-circular gears 12. 11, and the variation invelocity ratio R through the transmission over one revolution of theinput shafts 3.1. 12.1 when the second non-circular gears 12.11 areadvanced 110° relative to the first non-circular gears 3.4. With thisphase angle, the velocity ratio R of the transmission drops to 0.75 forthe period in which the overrunning clutch is locked. This intervalcoincides with the common interval in which the velocity ratios of thefirst and second non-circular gear pairs vary linearly so as to providea constant velocity ratio R.

Diagram 3 shows the variation in velocity ratios of the pair of firstnon-circular gears 3.4 and the pair of second non-circular gears 12.11,and the variation in velocity ratio R through the transmission over onerevolution of the input shafts 3.1. 12.1 when the second non-circulargears 12, 11 are delayed 110° relative to the first non-circular gears3.4, and the orientation of the overrunning clutch is reversed. Withthis phase angle, the velocity ratio R of the transmission increases to1.25 for the interval in which the overrunning clutch is locked. Again,this interval coincides with the common interval in which the velocityratios of the first and second non-circular gear pairs vary linearly soas to provide a constant velocity ratio R. Consequently, it will be seenthat the amplification of the “bevel-bevel” transmission can change by33% in one direction (between 110° and 0°), and 25% in the oppositedirection (between 0° and −110°).

FIG. 1 c shows another prior art infinitely-variable transmission, whichis substantially identical to the transmission shown in FIGS. 1, 1 a,except that the first non-circular gear 4 days splined to theintermediate shaft 13 and the cage 9, while the second bevel gear 10 iscoupled to the take-off gear 2. Consequently, the transmission shown inFIG. 1 c is referred to as a “bevel-carrier” transmission.

Diagrams 1, 2 and 3 of FIG. 1 d show the variation in velocity ratio W₄of the pair of first non-circular gears 3,4, the variation in velocityratio W₁₁ of the pair of second non-circular gears 12, 11, and thevariation in velocity ratio R through the transmission over onerevolution of the inputs shaft 3.1, 12.1. As will be apparent, thevelocity ratio of the first non-circular gears 3,4 increases linearlyover the first 220° of input shaft rotation, and then decreasesnon-linearly over the subsequent 140° of input shaft rotation.

The velocity ratio R is given by the equation R=2W₁₁−W₄. Consequently,the velocity ratio of the second non-circular gears 12, 11 alsoincreases linearly over the first 220° of input shaft rotation toprovide a constant velocity ratio R over this interval. The velocityratio of the second non-circular gears 12, 11 decreases over thesubsequent 140° of input shaft rotation.

As shown in Diagram 1, the velocity ratio R does remain constant at 1.0over the first 220° of input shaft rotation, as expected. Thereafter,the intermediate shaft 13 overruns the first bevel gear 8, causing thefirst and second non-circular gear pairs to become uncoupled from theinput and output shafts.

The velocity ratios R for phase angles of +110° and −110° are shown inDiagrams 2 and 3, respectively, of FIG. 1 d. From these diagrams it willbe apparent that the interval in which the overrunning clutch is lockedcoincides with the common interval in which the velocity ratios of thefirst and second non-circular gear pairs vary linearly so as to providea constant velocity ratio R. From these diagrams, it will also beapparent that the amplification of the “bevel carrier” transmissionshown in FIG. 1 c can change 50% in one direction and 100% in theopposite direction (with the orientation of the overrunning clutchreversed).

FIG. 1 e shows another prior art infinitely-variable “bevel-carrier”transmission. which is substantially identical to the “bevel carrier”transmission shown in FIG. 1c, except that the differential has a gearratio of 2:1. As shown in FIG. 1 f. the amplification of the“bevel-carrier” transmission shown in FIG. 1 e has changed 200% in onedirection with infinite translation in the opposite direction (with theorientation of the overrunning clutch reversed).

Turning now to FIG. 2, 2 a. one variable-ratio gear assembly of aninfinitely-variable transmission 100, according to a first embodiment ofthe present convention, is shown. As above. the transmission 100includes four identical variable-ratio gear assemblies which aredisposed preferably at equal angular intervals around the input shaft3.1 and the output shaft 1.1, although they may be disposed at unequalangular intervals with a possible reduction in performance. Forconvenience, only one variable-ratio gear assembly is shown.

The infinitely-variable transmission 100 comprises a “bevel-bevel”-typetransmission, similar in structure to the infinitely-variable“bevel-bevel” transmission shown in FIG. 1, 1 a. However, unlike the“bevel-bevel” transmission shown in FIG. 1, 1 a, the cage 5 of theone-way overrunning clutch is replaced with a sleeve 5.1 provided aroundthe intermediate shaft 13 and splined to the first non-circular drivengear 4. Also, each variable-ratio gear assembly of the transmission 100includes a programmable multi-directional coupling and an actuator, inreplacement of the one-way overrunning clutch.

The multi-directional coupling was described in published PCT PatentApplication No. 98/01072, and comprises a travelling inner conical race20 coupled to the sleeve 5.1 through a ball spline bx, a tubular outerclutch body 22 coupled to the first bevel gear 9.1, and a conical innerslipper 21 provided between the conical race 20 and the clutch body 22.The conical race 20 includes an outer conical friction surface, and theslipper 21 includes an inner conical friction surface which meets withthe conical friction surface of the conical race 20. The slipper 21includes an outer bearing surface. and the clutch body 22 includes aninner bearing surface which, together with the outer bearing surface ofthe slipper 21, defines a channel between the slipper 21 and the clutchbody 22. The channel includes a plurality of pockets for retainingroller elements 6 therein, disposed in abutment against the bearingsurfaces. Consequently when the conical race 20 is pressed axiallyinwards into the coupling while rotating relative to the clutch body 22,the outer friction surface of the race 20 engages in the inner frictionsurface of the slipper 21, causing the roller elements 6 to roll uptheir respective pockets and to press the slipper 21 inwardly againstthe conical race 20, thereby locking the conical race 20 to the clutchbody 22 in both directions. Conversely, when the conical race 20 ispulled axially from the coupling, the roller elements 6 roll down theirrespective pockets causing the slipper 21 to retract from the conicalrace 20, thereby allowing the conical race 20 to rotate relative to theclutch body 22 in both directions.

The actuator serves to insert and withdraw the conical race 20 from thecoupling, to thereby couple and uncouple the first non-circular gears3,4 and the second non-circular gears 12, 11 to and from the input andoutput shafts. The actuator comprises a first shadow disc cam 16 coupledto the first input shaft 3.1, a second shadow disc cam 16 b coupled tothe second input shaft 12.1. a shadow cam follower plate 17 rotatablydisposed around the sleeve 5.1 and including a plurality of aperturesextending therethrough, a plurality of double-ended decoupling cone pins19 disposed within the apertures, a stationary end ring 18 splined tothe sleeve 5.1 provided between the first non-circular driven gear 4 andthe shadow cam follower plate 17. a travelling end ring 24 splined tothe sleeve 5.1 provided between one end of the conical race 20 and theshadow cam follower plate 17, and a spring s1 provided adjacent theopposite end of the conical race 20. The shadow disc cams 16. 16 arotate against the shadow cam follower plate 17, each including a camlobe which is synchronized with the nonlinear portions of the velocityratios of the respective non-circular gear pairs 3.4; 12.11.

As the cam lobe rotates against the cam follower plate 17. the camfollower plate 17 is pushed upwards by the cam lobe, causing thetravelling end ring 24 to eject the conical race 20 from themulti-directional coupling and thereby uncouple the first bevel gear 8from the first non-circular driven gear 4. After the cam lobe rotatesaway from the cam follower plate 17, the cam follower plate 17 movesdownwards, causing the conical race 20 to be pressed into the couplingvia the force exerted by the spring s1.

Subdiagrams a) and b) of FIG. 2 b respectively depict the firstnon-circular gears 3, 4, and the corresponding shadow disc cam 16.Subdiagrams c) and d) of FIG. 2 b respectively depict the secondnon-circular gears 12.11, and the corresponding shadow disc cam 16 awhen the phase angle between the first and second non-circular gears is0°, +110° and −110. Subdiagram e) of FIG. 2 b shows that the commonangular period during which both of the first and second non-circulargears are coupled to the input and output shafts coincides with theangular interval in which the acceleration of both the first and secondnon-circular gear pairs is constant. As discussed above, the requirementensures that the velocity ratio R of the transmission remains constantover the period in which the first and second non-circular gear pairsare coupled to the input and output shafts. The interval during whichthe first and second non-circular gear pairs are uncoupled from theinput and output shafts occurs when the acceleration of either or bothof the first and second non-circular gears is nonuniform. Since theshadow disc cams 16. 16 a are coupled to the input shafts 3.1, 12.1,this latter angular interval coincides with the non-linear portions ofthe pitch circles of the drive gears 3, 12. As will be discussed below,the angular interval in which the first and second non-circular gearpairs are uncoupled from the input and output shafts may also coincidewith the non-linear portions of the pitch circles of the driven gears 4,11.

The graph shown in FIG. 2 b depicts the variation in velocity ratios ofthe pair of first non-circular gears 3,4 and the pair of secondnon-circular gears 12, 11, and the variation in velocity ratio R throughthe transmission 100 over one revolution of the inputs shaft 3.1, 12.1when the second non-circular gears 12, 11 lead the first non-circulargears 3,4 by 110°, and lag the first non-circular gears 3,4 by 110°. Aswill be apparent, the velocity ratio R through the transmission 100remains constant at 0.75 when the phase angle is +110°, and increases to1.25 when the phase angle is −110°. Consequently, the amplification ofthe transmission 100 is 166%, which represents a dramatic improvementover the prior art. Further, as the multi-directional couplings can lockand free-wheel in both directions, the transmission 100 can takeadvantage of engine braking. Therefore, the output shaft 1.1 may act asa torque input member, with the input shafts 3.1, 12.1 acting as torqueoutput members, if so desired. In this instance, the differential wouldact as a torque splitter which splits the input torque between the firstnon-circular gear pair and the second non-circular gear pair.

FIG. 2 c shows an infinitely-variable “bevel-carrier” transmission 200,according to a second embodiment of the present convention. Theinfinitely-variable transmission 200 is substantially identical to the“bevel-carrier” transmission shown in FIG. 1 c, except that the cage 5of the one-way overrunning clutch is replaced with a sleeve 5.1 providedaround the intermediate shaft 13 and splined to the first non-circulardriven gear 4, and the one-way overrunning clutch is replaced with aprogrammable multi-directional coupling and an cam-follower actuator. Asshown in FIG. 2 d, the amplification of the transmission shown in FIG. 2c has a changed to 300% in both directions, which again represents adramatic improvement over the prior art.

FIG. 2 e summarizes the non-circular gear pairs for the “bevel-bevel”transmission 100. and the “bevel-carrier” transmission 200. Diagrams a)and b) of FIG. 2 e depict the first and second non-circular gear pairsfor the “bevel-bevel” transmission 100. The first non-circular gearpairs 3.4 for the “bevel-bevel” transmission 100 are shown being of thesame size as the second non-circular gear pairs 12, 11, except that thesecond non-circular gears 12, 11 are flipped over and rotated 180° toprovide the velocity ratio profiles shown in FIG. 2 b. Diagrams b) andc) of FIG. 2 e depict the first and second non-circular gear pairs forthe “bevel-carrier” transmission 200. The first non-circular gear pairs3,4 for the “bevel-bevel” transmission 200 are oriented the same way asthe second non-circular gear pairs 12.11, but produce twice the angularacceleration as the second non-circular gears 12, 11 to provide thevelocity ratio profiles shown in FIG. 2 b.

Numerous variations of the foregoing embodiments may be realized. FIG. 2f shows an infinitely-variable “bevel-carrier” transmission 300,according to a third embodiment of the invention. The “bevel-carrier”transmission 300 is substantially identical to the “bevel-carrier”transmission shown in FIG. 1 e, except that the cage 5 of the one-wayoverrunning clutch is replaced with a sleeve 5.1 provided around theintermediate shaft 13 and splined to the first non-circular driven gear4. and the one-way overrunning clutch is replaced with a programmablemulti-directional coupling and an cam-follower actuator. Due to the 2:1gear ratio of the differential, the first and second non-circular gearpairs can produce the same angular acceleration. The amplificationcharacteristics of the transmission 300 are similar to those thetransmission shown in FIG. 1 e in so far as the transmission 300 iscapable of producing infinite translation. As will be discussed below,this feature is advantageous when the transmission 300 is used in anautomobile since it allows the output torque of the transmission 300 todrop to zero when the vehicle is stopped.

FIG. 2 g shows an infinitely-variable “bevel-carrier” transmission 400which is identical to the “bevel-carrier” transmission 200 shown in FIG.2 c, except that the differential is replaced with a planetary gear setcomprising an annulus 8, a pinion 10, a cage 9 and a planet 9.1. Withthe annulus/pinion ratio set equal to the ratio of the first and secondbevel gears 8, 10 of the differential, the amplification of thetransmission 400 will be the same as that for the transmission 200.

FIG. 2 h shows an infinitely-variable “bevel-carrier” transmission 500which is substantially identical to the “bevel-carrier” transmission 400shown in FIG. 2 g, except that the differential is replaced with acoplanar reverted geartrain loop. The coplanar reverted geartrain loopwas disclosed in published PCT Patent Application 98/01019, andcomprises a pinion 10, an annular internal gear 8 disposed around thepinion 10 and being coaxial to the pinion 10, and a cage assembly 9including a ring gear 9.1. The ring gear 9.1 includes an inner surfacewhich engages the pinion 10, and an outer surface which engages theangular gear 8. The cage assembly 9 also includes an eccentric guide fordisposing of a ring gear 9.1 coplanar to and eccentrically with respectto the pinion 10 and the angular gear 8. With the annulus/pinion ratioset equal to the ratio equal to 3/2, the amplification of thetransmission 500 will be the same as that for the transmission 200.

Thus far. all of the foregoing embodiments have employed a couplingincluding an inner conical race 20, as shown above and reproduced inFIG. 3 c. However, the invention is not so limited. Rather, the couplingmay include instead an outer conical race 20, a tubular inner clutchbody 22, and a conical slipper 21 provided between the outer conicalrace 20 and the inner clutch body 22. as shown in FIGS. 3 a, 3 b.Alternately, rather than the actuator inserting and withdrawing theconical race 20 from the coupling, the actuator may instead insert andwithdraw the slipper 21 from the coupling. as shown in FIGS. 3 b. 3 d.Numerous other variations of the coupling will be apparent to thoseskilled in the art.

Also, all of the actuators described thus far herein for coupling anduncoupling the first and second non-circular gear pairs to and from theinput and output shafts have comprised shadow disc cams 16, 16 a drivinga plurality of double-ended decoupling cones 19. However, other actuatorimplementations are possible. One such implementation is shown in FIGS.4 a-4 c, and comprises a pair of shadow disc cams 16, 16 a coupled torespective input shafts 3.1, 12.1. an annular shadow cam follower plate17 in abutment with the shadow disc cams 16, 16 a, a decoupler disc 57disposed within the cam follower plate 17 and including V-groovesprovided on both end faces of the decoupler disc 57, and roller bearingsb2 provided between the cam follower plate 17 and the decoupler disc 57for allowing the cam follower plate 17 to rotate relative to thedecoupler disc 57. The actuator also includes an axially-fixed annularcone race 56 provided between the first non-circular driven gear 4 andthe shadow cam follower plate 17 and including an axial cone mating withthe V-groove on one end face of the decoupler disc 57, and a travellingannular cone race 58 provided between the coupling and the shadow camfollower plate 17 and including an axial cone mating with the V-grooveon the opposite end face of the decoupler disc 57. Also, the travellingannular cone race 58 includes an axially-extending conical flangeopposite the axial cone portion thereof which is splined to the annulargear 8 through ball splines bx and replaces the conical race 20 of thecoupling.

The actuator shown in FIGS. 4 a-4 c operates in a similar fashion to theactuators described above. As the cam lobe rotates against the camfollower plate 17, the cam follower plate 17 is pushed upwards by thecam lobe, causing the travelling cone race 58 to move laterally awayfrom the cam follower plate 17, thereby uncoupling the annular gear 8from the second non-circular driven gear 11. After the cam lobe rotatesaway from the cam follower plate 17, the cam follower plate 17 movesdownwards, causing the travelling cone race 58 to move laterally towardsthe cam follower plate 17, thereby coupling the annular gear 8 to thesecond non-circular driven gear 11.

FIGS. 5 a-5 b depict another actuator implementation which replaces thecam follower plate 17 and the associated races 56, 58 with a rocker arm64, a tappet 60 and coil spring 63. Whenever a cam lobe presses upwardsagainst the rocker arm 64, the rocker arm 64 presses against the tappet60 thereby forcing the conical slipper 20 out of the coupling. When thecam lobe moves away from the rocker arm 64, the spring 63 releases thepressure on the conical slipper 20 by the tappet 60. causing the conicalslipper 20 to be drawn into the coupling. The advantage of thisvariation is that the rocker arm 64 provides a mechanical advantagewhich permits a shallower rise on the shadow disc cams.

FIGS. 6, 6 a depict an actuator similar to the actuator shown in FIGS. 4a-4 c, comprising a cam follower platen 68 disposed within the camfollower plate 17 and including V-grooves provided on one end face ofthe cam follower platen 68, and roller bearings provided between the camfollower plate 17 and the cam follower platen 68 for allowing the camfollower plate 17 to rotate relative to the cam follower platen 68. Theactuator also comprises an axially-fixed bearing end plate 70 providedbetween the first non-circular driven gear 4 and the shadow cam followerplate 17 and including a key 69 for preventing rotation of the camfollower platen 68, and a travelling bearing end plate 71 providedbetween one end of the conical race 20 and the shadow cam follower plate17 and including an axial cone mating with the V-groove on the camfollower platen 68.

The actuator shown in FIG. 6 b is similar to the actuator shown in FIGS.5 a-5 b, comprising a rocker arm 64′ and a platen 68′ which lie in thesame plane as the cam followers 16 a. 16 b. Also, the cam follower plate17 is eliminated in this variation. The platen 68′ includes ribs forengaging the conical race 20. Whenever a cam lobe presses outwardsagainst the rocker arm 64′. the rocker arm 64′ forces the platen 68′ tomove transversely to the axis of rotation of the intermediate shaft 13,thereby causing the ribs of the platen 68′ to engage the conical slipper20 and forcing the conical slipper 20 into the coupling. The advantageof this variation is that the actuator is not affected by centrifugalforces from rotation of the shadow disc cams.

All of the actuators described thus far herein have comprised shadowdisc cams 16, 16 a together with means for transferring the cam actionof the shadow disc cams to the coupling, and have shadowed thenon-linear portions of the pitch circles of the first and second drivinggears 3, 12. FIGS. 7 to 10 depict actuators employing shadow bearingcams for coupling and uncoupling the first and second non-circular gearpairs to the input and output shafts, and which shadow the non-linearportions of the pitch circles of the first and second driven gears 4,11. The actuator shown in FIGS. 7, 7 b, 7 c is shown being used inassociation with a “bevel-carrier” transmission which uses a coplanarreverted gear train loop 77 for transmitting torque between the firstand second non-circular gear sets, and the input and output shafts. Theannular gear is splined to the first driven gear 4, the pinion issplined to the second driven gear 11, and the cage assembly of thecoplanar reverted gear train loop 77 is coupled to the output gear 2through the coupling.

The actuator comprises a stationary platen cage 74 disposed around theintermediate shaft 13 and including a first and second thru-ports 74,75, a first movable ball follower 88 provided within the first thru-port74, and a second movable ball follower 88′ provided within the secondthru-port 75. The actuator also includes a first axial ball cam race 72coupled to the first driven gear 4 and being in communication with thefirst thru-port 74, a second axial ball cam race 73 coupled to thesecond driven gear 12 and being in communication with the secondthru-port 75, and an axial ball follower race 81 coupled to the conicalslipper 20 of the coupling and being in communication with thethru-ports 74, 75. The axial ball cam races 72, 73 are concentric toeach other. Also, each of the axial ball cam races 72, 73 includes araised race portion which is synchronized with the non-linear portion ofthe pitch circle of the respective driven gear for uncoupling the firstand second non-circular gears from the input and output shafts when theacceleration of either of the non-circular gear pairs is non-uniform.FIG. 7 a depicts the raised race portions of the axial ball cam races72, 73 for the respective non-circular gear pairs.

The actuator shown in FIGS. 8, 8 a is shown being used in associationwith a “bevel-bevel” transmission which uses a coplanar reverted geartrain loop 77 for transmitting torque between the first and secondnon-circular gear sets, and the input and output shafts. The pinion 10of the coplanar reverted gear train loop 77 is rotatably disposed arounda fixed countershaft 55 and is splined to the first driven gear 4. Thecage assembly 9.1 is splined to the second driven gear 11. The annulargear is coupled to the output gear 2 through the coupling.

The actuator comprises a first bearing half-race 83 provided on ashoulder of the pinion 10, and a second bearing half-race 82 disposedadjacent the radially innermost shoulder of the cage assembly 9.1. Thepinion 10 is rotatably disposed within the cage assembly 9.1 such thatthe first and second bearing half-races 82, 83 together comprise a fullbearing race. The actuator also includes a single ball follower 88provided within the full bearing race. and a stationary ball followerguide 74 provided between ball follower 88 and the conical race 20 ofthe coupling. Also, the first and second bearing half-races 82. 83 eachinclude a raised race portion which is synchronized with the non-linearportion of the pitch circle of the respective driven gear for uncouplingthe first and second non-circular gears from the input and output shaftswhen the acceleration of either of the non-circular gear pairs isnon-uniform.

The actuator shown in FIGS. 9, 9 a is similar to the actuator shown inFIG. 8, 8 a. comprising a first bearing half-race 80 and a secondbearing half-race 82 concentric with the first bearing half-race 80,both provided adjacent the radially innermost shoulder of the cageassembly 9.1. The actuator also comprises a ball cam riser 83 includinga first riser half-race 83′ and a second riser half-race 83″ concentricwith the first riser half-race 83′. The ball cam riser 83 is retained ina channel within the cage assembly 9.1, and includes a key which extendsthrough the cage assembly 9.1 and engages a mating key on the secondbevel gear 10 to allow the ball cam riser 89 to rotate with the secondbevel gear 10. The first bearing half-race 80 and the first riserhalf-race 83′ together comprise a first axial ball cam race. Similarly,the second bearing half-race 82 and the second riser half-race 83″together comprise a second axial ball cam race. The actuator alsoincludes a stationary ball follower platen 74 including a pair ofthru-ports for receiving a movable ball follower therein, and a followerball race plate 81. The stationary ball follower platen 74 is keyed tothe fixed countershaft 55. The follower ball race plate 81 is coupled tothe outer conical slipper of the coupling, and includes a pair of ballfollower races coinciding with the first and second axial ball camraces. Also, the first plate half-race 80 and the second riser half-race83″ each include a raised race portion which is synchronized with thenon-linear portion of the pitch circle of the respective driven gear foruncoupling the driven gears 4, 11 from the input and output shafts whenthe acceleration of either of the non-circular gear pairs isnon-uniform.

The actuator shown in FIGS. 10, 10 a is shown being used in associationwith a “bevel-carrier” transmission which uses a coplanar reverted geartrain loop for transmitting torque between the first and secondnon-circular gear sets, and the input and output shafts. The pinion ofthe coplanar reverted gear train loop is splined to the first drivengear 4. The cage assembly is fixed to the intermediate shaft 13. Theannular gear is coupled to the second driven gear 11 through thecoupling.

The actuator comprises a first centre axis gear 85 coupled to the firstdriving gear 3. a second centre axis gear 84 coupled to the seconddriving gear 12, a first intermediate gear 85′ rotatably disposed aroundthe intermediate shaft 13 and meshing with the first centre axis gear85, a second intermediate gear 84′ rotatably disposed around theintermediate shaft 13 and meshing with the second centre axis gear 84,and a ball follower guide 86 disposed around the intermediate axis 13between the intermediate gears 84′, 85′ and the conical slipper of thecoupling. The ball follower guide 86 includes a pair of apertures forreceiving truncated ball followers 88 therein. Also, the first andsecond intermediate gears 84′, 85′ each include a respectiveaxially-extending riser portion 91, 90 which is synchronized with thenon-linear portion of the pitch circle of the respective driving gearfor uncoupling the first and second non-circular gears from the inputand output shafts when the acceleration of either of the non-circulargear pairs is non-uniform.

Having described various implementations of the variable rationon-circular gear pairs, the multi-directional coupling and theactuator, an implementation of the phase angle variator for rotating thesecond non-circular gear pair 12.11 relative to the first non-circulargear pair 3, 4 will now be discussed with reference to FIG. 11. Thephase-angle variator shown in FIG. 11 is a hydraulically-operatedvariator, and comprises a stator 25 splined to the first input shaft3.1. and a rotor 26 provided within the stator 25. The rotor 26 rotatesabout an axis which is eccentric to the centre of the stator 25. andincludes a lobe extending radially outwards from the rotor body to theinner surface of the stator 25. The rotor 26 includes a cylindricalaxial extension 26.1 splined to the second input shaft 12.1incorporating an aperture p1 for receiving pressurized hydraulic fluid.and a sleeve 28 pressure fitted within the rotor 26. The phase anglevariator also includes an oil spool 27 splined to the first input shaft3.1 and being positioned inside the rotor 26 within the sleeve 28. Theoil spool 27 includes a plurality of spiral fluid passageways providedbetween lands 27.1, 27.2 disposed around the circumference of the oilspool 27. The sleeve 28 includes fluid ingress and egress ports +, − forpassing pressurized fluid p2 to and from the fluid passageways. Thephase angle variator also includes a plurality of springs s2 providedagainst one end of the oil spool 27 for urging the oil spool 27 into therotor 26.

In operation, pressurized hydraulic fluid is applied to the aperture p1of the rotor 26, thereby urging the oil spool 27 to move axiallyrelative to the rotor 26 in opposition to the force exerted on the oilspool 27 by the springs s2. As the oil spool 27 moves axially, the lands27.1, 27.2 are displaced a distance from the fluid ingress and egressports +, − thereby allowing fluid to enter the fluid passageways. Thepressurized fluid in the fluid passageways causes the rotor 26 (and thefirst input shaft 3.1) to rotate relative to the second input shaft 12.1until such time as the lands 27.1, 27.2 are rotated into a positionwhich seals off the fluid ingress and egress ports +, −. As will beappreciated. the direction of rotation of the rotor 26 depends on thewhether the pressurized fluid p2 is applied to the fluid port + or thefluid port

Diagram 6 of FIG. 11 shows an “unrolled” oil spool 27, depicting thelands 27.1, 27.2, and shows the angular displacement of the oil spool 27relative to the pressure of the hydraulic fluid at the port p1. As willbe apparent, by controlling the pressure of the hydraulic fluid at theport p1, the phase angle between the first and second input shafts 3.1,12.1 can be varied.

As discussed above, infinite translation is possible with thetransmissions according to the present invention. For instance, theinfinitely-variable “bevel-carrier” transmission 300 shown in FIG. 2 fwould have infinite translation when the phase angle is 110°. However,if the phase angle was not exactly 110°, and the transmission was fittedinto an automobile, the automobile would be subject to creep. Diagram 7of FIG. 11 shows a variation to the lands 27.1, 27.2 which is intendedto prevent vehicle creep. The lands 27.1, 27.2 are modified by includingrespective non-spiral portions which are designed to seal off the fluidingress and egress ports +, − when the phase angle is betweenapproximately 90° and 120°. Consequently, within this angular range, thephase angle variator is disabled, and the reverse torque applied to therotor 26 will cause the rotor 26 to rotate until the output of thetransmission is zero.

FIG. 11 a shows another hydraulically-operated variator, similar to thevariator shown in FIG. 11, comprising a stator 25 splined to the firstinput shaft 3.1, and a rotor 26 provided within the stator 25 and beingcoupled to the second input shaft 12.1. However, unlike the variatorshown in FIG. 11, the variator shown in FIG. 11 a includes two lobesextending radially outwards from the rotor body to the inner surface ofthe stator 25 for defining four pressurization chambers around the rotor26 to obtain greater torque on the rotor 26. Also, the variator includesa pinion 30, and the rotor 26 includes an annular gear 26.1 coaxial withthe axis of rotation of the rotor 26 to obtain greater torque forrotation of the second input shaft 12.1. FIG. 11 b shows anotherhydraulically-operated variator similar to the variator shown in FIG. 11a, except that the rotor 26 includes a cage assembly incorporating aring gear 32 and an eccentric guide for meshing the outer gear surfaceof the ring gear 32 with the annular gear 26.1 and for meshing the innergear surface of the ring gear 32 with the pinion 30. As will beapparent. the latter rotor structure resembles a coplanar reverted geartrain loop. and eliminates the offset between the centre of the stator25 and the axis of rotation of the rotor 26.

FIG. 12 shows a phase-angle variator which varies the phase anglebetween the first and second variable ratio gear pairs in accordancewith the torque present at the output shaft 1.1. The variator comprisesthe first input shaft 3.1, the second input shaft 12.1. the firstnon-circular driven gear 3 splined to the first input shaft 3.1, and thesecond non-circular driven gear 12 splined to the second input shaft12.1. The variator also includes a torsion spring housing 93 splined tothe first input shaft 3.1, a torsion spring 92 disposed within thetorsion spring housing 93 and being coupled between the first inputshaft 3.1 at one end and the second input shaft 12.1 at the other end,and a stop 93.1 provided within the housing for limiting the magnitudeof the phase angle between the first and second variable ratio gearpairs. In operation, with a prime mover rotating the first input shaft3.1 with a torque t1, as shown in FIG. 12, and a load is applied to theoutput shaft 1.1, the second input shaft 12.1 will experience a reversetorque t2 which is proportional to the load applied. Consequently, theangle between the first and second variable ratio gear pairs will varyin accordance with the applied load and the spring constant of thetorsion spring 92. As will be appreciated. as the applied load increase,preferably the second input shaft 12.1 rotates relative to the firstinput shaft 3.1 sufficiently so as to reduce the torque on the primemover to an acceptable level.

It should also be pointed out that the width of the torsion springhousing 93 narrows with the radial distance from the input shafts. It isbelieved that this shape will bind the radially outermost portions ofthe torsion spring 92 to the torsion spring housing 93 so as to providegreater control over the phase angle.

FIGS. 13, 13 a, 13 b show a manually-operated variator which is suitablefor use on farm vehicles, such as tractors. The transmission shown inFIG. 13 includes the “bevel-carrier” variable-ratio gear assemblydescribed with reference to FIG. 2 h, and the actuator described withreference to FIG. 10. The variator comprises a variator differential,comprising a bevel gear control arm 96 splined to the first non-circulardriving gear 3, a variator bevel gear 97 splined to the secondnon-circular driving gear 12, and a cage 99 including a plurality ofpinions 98 meshing with the bevel gear control arm 96 and the variatorbevel gear 97. As will be appreciated, by driving the first input shaft3.1 with a prime mover, the velocity ratio of the transmission can becontrolled by rotating the control arm 96. Preferably, the magnitude ofthe phase angle is limited to 110° to avoid undesirable fluctuations inrotational output speed.

FIG. 14 shows an all-gear variator comprising a first control gear asplined to the first driven gear 4, a second control gear b splined tothe second driven gear 11, a cage assembly 807 rotatably disposed aboutthe input shafts 3.1, 12.1, and a spool 808 provided within the cageassembly 807 and retaining a gear pair 809 rotatable thereon. The gearpair 809 includes a first spool gear b meshing with the first controlgear a, and a second spool gear c fixed to the first spool gear b andmeshing with the second control gear d.

The variator includes a band clutch provided between the cage assembly807 and the casing 15 of the transmission which acts as a brake forselectively preventing rotation of the cage assembly 807 relative to thecasing 15. The band clutch comprises a clutch stator 801, a clutch rotor803 provided within the clutch stator 801, end plates 802 a. 802 b. anda clutch band 804 provided within the clutch rotor 803 and being coupledbetween the clutch rotor 803 at one end and the clutch end plates 802 atthe opposite end. The variator also includes a variator end plate 801splined to the second input shaft 12.1. and a cone coupling 806 andassociated cone actuator 805 for selectively preventing rotation of thecage assembly 807 relative to the second input shaft 12.1.

As will be apparent from FIG. 14, the diameter of the first control geara is greater than the diameter of the second control gear d.Consequently, if the brake is activated and the input shafts 3.1, 12.1are rotating in the same direction, a torque will be developed at thespool 808 while will force the second input shaft 12.1 to rotate aheadof the first input shaft 3.1, thereby increasing the phase angle. On theother hand, if the brake is released, the torque at the spool 808 willcause the cage assembly 807 to rotate, thereby forcing the second inputshaft to move back towards the first input shaft 3.1, thereby decreasingthe phase angle. Once the desired phase angle is reached. the conecoupling 806 is activated and the brake is released (if not alreadyreleased), thereby preventing further movement of the second input shaft12.1 relative to the first input shaft 3.1.

Thus far, the foregoing discussion has assumed that the non-circulargears 3.4, 11, 12 have used standard involute-shaped gear teeth.Although the such gear teeth would be acceptable for use in theinfinitely-variable transmissions of the present invention, thevariations in the pitch circle diameters would give rise to variationsin contact ratio between the driving non-circular gears 3.12 and theassociated driven non-circular gear 4.11. Consequently the load whichcould be carried by the transmissions would be limited. Further, thevariations in contact ratio would cause excessive gear noise. Therefore,it is desirable to have non-circular gears having constant contactratio.

FIG. 15 depicts the conventional method for generating involute gearteeth on circular gears. First, a pitch circle pc is defined for eachgear corresponding to the gear ratio of the gear pair. Then, a pressureangle is then selected, and a generating circle gc is constructed foreach non-circular gear such that the line of action passes through thepitch point a of the pitch circles and is tangential to each pitchcircle gc. Since the line of action represents a line of congruency, allpoints on the tooth flank must lie on the line of action as the toothrotates towards the pitch point. Therefore, to determine the startingpoint on the tooth flank a line m is drawn tangential to the line ofaction and passing through the pitch point. This line m is then rotatedabout a line n tangential to the line m and passing through therespective centre of each gear until the pitch point intersects thecorresponding pitch circle. This point of intersection is denoted as a′.

The point of intersection of the line m with the generating circle isdenoted as 1. As shown in Drawing 1 of FIG. 15, angle θ denotes theangular displacement of the point a′ relative to the pitch point a. Thepoint 1 is then rotated about the respective gear centre towards thepitch point over the angle θ. The resulting point represents thestarting point for the tooth flank, and is denoted in Diagram as 1″ forgear 3 and 1′ for gear 4. This process is then repeated for each pointon-the line of action between the starting points 1 for each gear, withthe resulting locus defining the shape for each tooth flank. Exampleloci are denoted in Diagram 1 as x and y. However, this process isunsuitable for non-circular gear pairs since the pitch point a movesalong the line between the gear centres as the gears rotate.

The following method has been developed for defining tooth flanks fornon-circular gears while ensuring that the gears have a constant contactratio. Although the following method is directed to non-circular gearswhen used in conjunction with infinitely-variable transmissions. itshould be understood that the method is applicable to any set of gearsin which a constant contact ratio is desired.

First, the pitch locus for the gears must be defined in accordance withthe desired velocity ratio profile. FIG. 15 a depicts the velocity ratioprofile R for a first non-circular gear pair 3, 4 (shown in FIG. 15 bfor convenience). As will be apparent. the average velocity ratio of thegear pair 3, 4 is 1.0 since the non-circular gear 3 has the same size asthe non-circular gear 4. However, the following discussion is alsorelevant to producing gears of unequal size.

From the velocity ratio profile shown in FIG. 15 a, it is possible toobtain the polar coordinates for each point on the pitch locus for thedriving gear 3. For instance, when the driving gear 3 has rotated 110°,the velocity ratio R is 1.0. Consequently, this point on the drivinggear 3 will lie 50% of the distance between the centre of the drivinggear 3 and the centre of the driven gear 4 and, therefore, will have asits polar coordinates 0∠110°.

Also, from this profile, it is possible to obtain the polar coordinatesfor each point on the pitch locus for the driven gear 4 by determiningthe area under the velocity ratio profile. For instance, the point onthe driven gear 4 which will provide a velocity ratio R of 1.0 must alsohave a radius of 0.5 relative to the distance between the centre of thedriving gear 3 and the centre of the driven gear 4. Since the velocityratio of this point will start at 0.5 (at 0°) and increase linearly to1.0 (at 110°), this point will have an average velocity ratio of 0.75over 110°. Consequently, the angular displacement of this point will be82.5° (75% of 110°), and its polar coordinates will be 0.5∠82.5°. Thepolar coordinates for the non-linear portion of the velocity ratioprofile can be determined by integrating the equation for the velocityratio profile over the angular interval of the non-linear portion. FIG.15 c depicts the polar coordinates for each point on the pitch locus forthe non-circular gears 3, 4.

FIG. 15 d depicts the velocity ratio profile R for a second non-circulargear pair 12, 11 (shown in FIG. 15 e for convenience). As will beapparent, the second non-circular gears 12, 11 provide less accelerationthan the first non-circular gears 3,4. FIG. 15 f depicts the polarcoordinates for each point on the pitch locus for the non-circular gears12, 11.

FIG. 15 g depicts a variation of the velocity ratio profile R, shown inFIG. 15 a, 15 d, characterized by an upward irregularity in the rotationof a driven variable ratio gear near the end of a period of constantacceleration, or by a downward irregularity in the rotation of a drivenvariable ratio gear near the beginning of a period of constantacceleration, and/or vice versa. These irregularities are denoted by +aand −a in FIG. 15 g. This variation provides a slight slackening off intorque during a coupling phase in a torque continuum. therebyfacilitating uncoupling of an active programmable coupling and,therefore, transfer of torque between angularly-adjacent variablevelocity-ratio gear sets.

After the pitch locus for each non-circular gear has been defined,preferably the pitch locus is divided into equal angular segmentscorresponding to the number of teeth desired so that constant contactratio is obtained. The pitch locus shown in FIG. 15 h has been dividedinto 12° segments to provide 30 teeth on each non-circular gear. Then, asegment of the pitch locus is selected, and points are plotted along onehalf of the segment beginning with the point on the pitch locus whichintersects the line joining the gear centres. In Diagram 1 of FIG. 15 h,these locus points are denoted as a, b, c, d e, f, g. Each locus point athru g is then projected back to the line joining the centres of thegears. thereby defining the effective pitch point for each pitch locuspoint. The effective pitch points are denoted as pb, pc, pd, pe, pf, pgin Diagram 1.

Once the effective pitch points are defined for the pitch locus points.an effective line of action is created for each locus point, passingthrough the effective pitch point. Since all the locus points aredesigned to lie on the actual line of action of the gears. the effectivelines of action are drawn parallel to each other and inclined with thesame angle as the actual line of action. In Diagram 1 of FIG. 15 h. theline of action is 25°. as in the case of the circular gear set of FIG.15. Then, each pitch locus point is projected back along the respectiveeffective line of action to locate an effective generating circle point.The distance between the effective generating circle point to therespective effective pitch point is equal to the arc length between therespective pitch locus point and the line joining the gear centres. Theeffective generating circle points are denoted as ab, ac, ad, ae, af, agin Diagram 1.

Finally, the shape of the tooth flank is defined by rotating eachgenerating circle point over an angular interval θ equal to the angularinterval between the respective pitch locus point and the line joiningthe gear centres. By rotating each generating circle point about thecentre of each gear, the addendum of the tooth flank of one gear and thededendum of the tooth flank of the opposite gear will be created. Theprocess is then repeated for the locus points on the other half of thepitch locus segment. This method ensures that the tooth flank of onegear remains congruent with the tooth flank of the meshing gear over anangular span which is twice the angular interval between similar toothflanks. Therefore, a contact ratio of at least 2.0 is attained.Alternately, as shown in FIG. 15 i, the process may be conducted for theentire tooth flank in a single step by projecting each pitch locus pointalong the respective effective line of action a distance equal to halfthe arc length between the respective pitch locus point and the linejoining the gear centres. and then rotating the effective generatingcircle points back by angular interval θ/2.

Having described several embodiments of the invention, a number ofpreferred transmission implementations incorporating these embodimentswill now be briefly described. FIG. 16 depicts a transmission whichincorporates the hydraulically-controlled variator described withreference to FIG. 11, 11 a, 11 b; the infinitely-variable“bevel-carrier” transmission 500; and Hi-Lo-Reverse gear box coupled tothe output of the transmission 500.

FIG. 16 a depicts a transmission which is substantially similar to thetransmission shown in FIG. 16, but including a spool release clutchassembly 35 which allows the casing 15 to rotate in unison with theinput and output shafts when the velocity ratio of the transmission 500reaches unity. The casing 15 is shown in more detail in FIG. 16 b. Thismodification to the transmission shown in FIG. 16 is advantageous sincethe gears of the transmission 500 will function as a solid coupling whenthe velocity ratio reaches unity. Therefore, by coupling the casing 15to the input and output shaft at unity velocity ratio reduces power lossat high speeds by effectively removing the transmission 500 from thetorque continuum.

FIG. 17 depicts a transmission which incorporates an infinitely-variabletransmission which amalgamates the features described with reference toFIGS. 9 and 10; the phase-angle variation described with reference toFIG. 12; and a 2:1 gear box coupled to the output of theinfinitely-variable transmission so as to obtain gear ratios extendingin infinitely-small increments from 1:1 to 4:1.

The foregoing embodiments are intended to be illustrative of thepreferred embodiments of the invention. Those of ordinary skill mayenvisage certain additions, deletions or modifications to the foregoingembodiments which, although not specifically suggested herein. will notdepart from the spirit or scope of the invention as defined by theappended claims.

1-7. Canceled
 8. An infinitely-variable transmission comprising: arotational input member and a rotational output member: a pair ofvariable velocity-ratio gear sets, a multi-directional couplingassociated with the gear sets, an actuator associated with the couplingfor coupling the gear sets to the rotational members over a commonangular period for providing a uniform velocity ratio between therotational members over the angular period, a phase angle variatorassociated with at least one of the gear sets for varying a rotationalangular displacement between the gear sets for varying the uniformvelocity ratio, and wherein each said variable-ratio gear set providesan interval of uniform acceleration and an interval of non-uniformacceleration, and the angular period comprises an angular interval ofthe uniform acceleration common to both of the gear sets.
 9. Thetransmission according to claim 8, wherein each said variable-ratio gearset comprises a pair of meshing non-circular gears, each saidnon-circular gear including a linear pitch circle portion and anon-linear pitch circle portion, and the angular period coincides withan interval of the linear pitch circle portions common to both of thegear sets.
 10. The transmission according to claim 8, wherein the phaseangle variator comprises a stator coupled to one of rotational members,and a rotor provided within the stator and being coupled to the other ofthe rotational members.
 11. The transmission according to claim 10,wherein the stator includes an aperture for receiving pressurized fluidfor displacing the rotor axially within the stator, and a plurality ofspiral passages, and the rotor includes pressurized fluid ingress andegress ports for communication with the spiral passages for rotating therotor as the rotor is axially displaced.
 12. The transmission accordingto claim 8, wherein the phase angle variator comprises a spring coupledbetween the rotational members for varying the angular displacement inresponse to output load.
 13. The transmission according to claim 8,wherein the phase angle variator comprises a differential including afirst bevel gear coupled to one of the rotational members, amanually-operated second bevel gear coupled to the other of therotational members, and a pinion coupled to the bevel gears.
 14. Thetransmission according to claim 8, wherein the phase angle variatorcomprises a first actuator gear coupled to one of the rotationalmembers, a second actuator gear coupled to the other of the rotationalmembers, a cage rotatable about the actuator gears and including a spoolrotatably coupled to the cage, the spool retaining a first spool gearthereon meshing with the first actuator gear, a second spool gearthereon meshing with the second actuator gear, the first spool gearbeing coupled to the second spool gear for rotation therewith and havinga diameter different from that of the second spool gear, and a brake forselectively preventing rotation of the cage.
 15. The transmissionaccording to claim 14, wherein the phase angle variator furthercomprises a clutch coupled to the cage for selectively preventingrotation of the cage relative to the rotational members. 16-24. Canceled25. An all-gear transmission comprising: a rotational input member, anda rotational output member; a pair of variable velocity-ratio gear sets,the velocity-ratios being phased through a rotational phase angle; amulti-directional coupling associated with the gear sets; and anactuator associated with the coupling for coupling the gear sets to therotational members over a common angular period for providing a uniformvelocity ratio between the rotational members over the angular period,the uniform velocity ratio being dependent upon the rotational phaseangle.
 26. An infinitely-variable transmission comprising: a rotationalinput member, a rotational output member; a pair of non-circular drivinggears coupled to one of the rotational members; and a plurality ofvariable velocity-ratio gear assemblies disposed about the onerotational member, and being coupled to the non-circular gears and theother of the rotational members for providing a uniform velocity ratiobetween the rotational members over a rotational period, each saidvariable velocity-ratio gear assembly comprising: an intermediate shaftincluding a pair of non-circular driven gears meshing with thenon-circular driving gears, one of the driving gears and the associateddriven gear comprising a first variable-ratio gear pair, and the otherof the driving gears and the associated driven gear comprising a secondvariable-ration gear pair; a multidirectional coupling associated withthe variable-ratio gear pairs; and an actuator associated with thecoupling for coupling the variable-ratio gear pairs to the rotationalmembers over an angular period for providing the uniform velocity ratiobetween the rotational members over the angular period, the angularperiods for the variable velocity-ratio gear assemblies togethercomprising the rotational period; and a phase angle variator associatedwith the variable-ratio gear pairs for varying a rotational angulardisplacement between the first gear pairs and the second gear pairs forvarying the uniform velocity ratio.
 27. An actuator for transmittingpower between a pair of rotational drive members over an angular portionof a revolution of one of the drive members, a first of the drivemembers including a drive element and a second of the drive membersincluding a driven element, the actuator comprising: an intermediaterotational member, a first intermediate drive element meshing with thedriven element; a second intermediate drive element meshing with thedriven element, one of the intermediate drive elements being rotatablycoupled to the intermediate member and the other of the intermediatedrive elements being fixed to the intermediate member; a couplingcoupled between the intermediate member and the one intermediate driveelement; and a cam coupled to the one drive member and the coupling foraltering a coupling state of the coupling.
 28. The actuator according toclaim 27, wherein the multi-directional coupling comprises: a raceincluding a first tubular friction surface, a tubular member including afirst bearing surface, a tubular slipper including a second tubularfriction for coupling to the first tubular friction surface, and asecond bearing surface opposite the second friction surface, the secondbearing surface being coaxial to the first bearing surface and, togetherwith the first bearing surface, defining a channel disposedtherebetween, and a plurality of roller elements disposed in the channelin abutment against the bearing surfaces, the channel including a pocketretaining at least one of the roller elements therein for coupling therace to the tubular member as the tubular member and the slipper rotaterelative to one another.
 29. The actuator according to claim 28, whereinthe first friction surface comprises a conical friction surface, thesecond tubular friction surface being shaped to mate with the conicalfriction surface, and the cam comprises a bearing race, and a bearing incommunicating with the bearing race, the bearing race including a raceportion for pressing the coupling race against the slipper over theangular portion.
 30. The actuator according to claim 28, wherein thefirst friction surface comprises a conical friction surface, the secondtubular friction surface being shaped to mate with the conical frictionsurface, and the cam comprises a bearing race, and a bearing incommunication with the bearing race, the bearing race including a raceportion for pressing the coupling race against the slipper over theangular portion.
 31. The actuator according to claim 28, wherein thefirst friction surface comprises a conical friction surface, the secondtubular friction surface being shaped to mate with the conical frictionsurface, and the cam comprises a gear including an axially-extendingriser portion for pressing the race against the slipper over the angularportion.
 32. A method for defining tooth flanks on pairs of meshingnon-circular gears, comprising the steps of: determining a pitch locusfor one of the non-circular gears; segmenting the pitch locus into pitchlocus portions; determining an effective pitch circle locus for thepitch locus portions by projecting the pitch locus portions onto acenter line joining centers of the non-circular gears; determining aneffective generating circle locus for the effective pitch circle locus,the effective generating circle locus being determined in accordancewith a desired pressure angle between the non-circular gears; anddetermining a locus of congruency for the gears from the effectivegenerating circle locus.
 33. The method according to claim 32, whereinthe step of determining the effective pitch circle locus comprisesdefining a plurality of pitch circle points on the pitch locus portion,and rotating the pitch circle points about the center of the onenon-circular gear.
 34. The method according to claim 33, wherein thestep of determining the effective generating circle locus comprisesdetermining an arc length distance along the pitch locus portion betweenthe center line and each said defined pitch circle point, and projectingthe respective rotated pitch circle points the respective arc lengthdistance along a respective line in accordance with a desired line ofaction between the gears.
 35. The method according to claim 34, whereinthe step of determining the locus of congruency comprises determining anangular displacement of the defined pitch circle points relative to thecenter line, and rotating the projected pitch circle points therespective angular displacements about the centers of the gears.